Refrigeration process, apparatus and method

ABSTRACT

A method in which a fixed size orifice is selected and inserted between the evaporator outlet and the suction inlet of a compressor in a refrigeration system to restrict gas flow to the compressor and thereby decrease the peak torque required to operate the compressor. The disclosed apparatus is a hermetic compressor unit with such a calibrated orifice installed in each of the compressor intake tubes, thereby allowing the use of an electric drive motor which otherwise would be incapable of developing sufficient peak load torque after starting to operate the compressor in a refrigeration system with an evaporator at ambient temperature in the absence of such orifices.

United States Patent 11 1 1111 3,763,69 Hover ct. 9, 1973 [5REFRIGERATION PROCESS, APPARATUS 2,737,030 3/1956 Philipp 62/508 x ANDMETHOD 3,401,873 9/1968 Privon et a]. 417/312 x [75] Inventor: Paul B.Hover, Clinton, Mich. [73] Assignee: Tecumseh Products Company,

Tecumseh, Mich.

[22] Filed: Feb. 2, 1972 [21] Appl. No.: 222,733

[52] US. Cl 62/115, 62/215, 62/217, 62/296, 62/505, 62/508, 62/511,417/299 [51] Int. Cl. FZSb 41/06 [58] Field of Search 62/83, 115, 215,62/217, 296, 508, 511, 505, 527; 417/299, 312

[56] References Cited UNITED STATES PATENTS 3,264,842 8/1966 Dobbie62/508 X 2,133,875 10/1938 Steenstrup 417/312 2,198,258 -4/l940 Money62/296 2,445,527 7/1948 Hirsch 62/296 X 2,497,668 2/l950 Grumblatt62/508 X Primary Examiner-William F. ODea Assistant Examiner-Peter D.Ferguson Attorney-Arthur Raisch et a1.

[5 7] ABSTRACT A method in which a fixed size orifice is selected andinserted between the evaporator outlet and the suction inlet of acompressor in a refrigeration system to restrict gas flow to thecompressor and thereby decrease the peak torque required to operate thecompressor. The disclosed apparatus is a hermetic compressor unit withsuch a calibrated orifice installed in each of the compressor intaketubes, thereby allowing the use of an electric drive motor whichotherwise would be incapable of developing sufficient peak load torqueafter starting to operate the compressor in a refrigeration system withan evaporator at ambient temperature in the absence of such orifices.

27 Claims, 5 Drawing Figures PATENTED 91973 SHEET 2 [IF 2 This inventionrelates to refrigeration systems and more particularly to a vaporcompression refrigeration process, apparatus and method of constructingapparatus utilizing the process.

A vapor-compression refrigeration system with a hermetically sealedcompressor and a refrigerant such as R-12 is used in most conventionalhome refrigerators and freezers. Under normal operating conditions insuch a conventional refrigerating system, the gaseous refrigerant isdischarged from the evaporator and received at the suction inlet of thecompressor at a pressure in the neighborhood of pounds per square inchabsolute (p.s.i.a.). However, if the evaporator becomes abnormallyheated, such as by defrosting of the refrigerator or freezer or becausethe system is at ambient temperature and is being put into operation forthe first time or after a prolonged shutdown, the pressure of therefrigerant at the suction inlet to the compressor may increaseapproximately three-to-five fold. During a prolonged shutdown of sayover 12 hours, the pressure and temperature tends to become equalizedthroughout the refrigeration system, such equalization being referred toas temperature equilibrium or equalization. During a short shutdown, thepressure tends to become equalized while the temperatures are still notequalized, such equalization being referred to as pressure equalization.

When the compressor is first started in either a pressure equalized ortemperature equalized refrigeration system under such-abnormally hightemperature conditions, there is an abnormally large mass flow rate ofrefrigerant through the compressor and into the condenser which resultsin a temporarily abnormally high pressure at the discharge of thecompressor. After the refrigeration system has operated for a fewminutes, the mass flow rate of refrigerant through the compressor andthe discharge pressure return to normal. However, the temporaryabnormally large mass flow rate of refrigerant through the compressorand the resulting high discharge pressure require considerably greatertorque or work to run the compressor. Hence, such a refrigeration systemhas hitherto required an electric motor or other prime mover capable ofproducing the abnormally high torque required to run the compressorunder such temporary abnormal load conditions.

In the method of this invention the abnormally high mass flow rate ofrefrigerant through the compressor 7 during start-up of a pressureand/or temperature equalized refrigeration system is substantiallydecreased by coupling the discharge of the evaporator to the inlet ofthe compressor through an inexpensive orifice of a preselected fixedsize. The term orifice as referred to here and throughout the remainderof this specification is intended to mean one or more nozzle orsharp-edged orifices (including thin plate orifices), with pluralorifices being arranged in parallel. The orifice automatically throttlesabnormally high mass fiow rates of the refrigerant so that thecompressor is not unduly loaded during this temporary load condition andyet allows a sufficient mass of refrigerant to flow into the compressorduring normal operation of the refrigeration system. The use of theorifice thus allows the compressor to be run by a less powerful electricmotor without materially adversely affecting the overall efficiency ofthe refrigeration system. Such an electric motor is less expensive tomanufacture and requires less space in a hermetic unit than wouldotherwise be required. Apparatus constructed pursuant to the method ofthis invention includes a calibratedorifice of a given selected size.coupled to the inlet of a conventional compressor gas pump so as toenable it to be driven by an electric motor of a selected reducedcapacity; i.e., a motor having a maximum torque rating below thatotherwise required to initially start-up and drive the compressor in arefrigeration system with an abnormally large mass flow rate conditionwere the orifice not present, as described in greater detail hereafter.

Object of this invention are to provide a method of constructing acompressor for a mechanical vapor compression refrigeration system whichsubstantially decreases the torque load imposed on a compressor in apressure and/or temperature equalized refrigeration system, and amotor-driven compressor unit constructed pursuant to the method for usein such a refrigeration system which is compact and of economical andreliable construction.

These and other objects, features and advantages of this invention willbe apparent from the following description, claims and accompanyingdrawings in which:

FIG. 1 is an end view partially in vertical section of a hermeticallysealed compressor unit with an electric drive motor for use in a vaporcompression refrigeration system embodying this invention.

FIGS. 2 and 3 are side and end views, respectively, of a restrictedorifice of the hermetically sealed compressor unit of FIG. 1.

FIG. 4 is an isometric view partially in section of the compressor unitof FIG. 1 with the hermetic shell and various other component partsremoved therefrom.

FIG. 5 is a semi-schematic drawing of a vapor compression refrigerationsystem embodying this invention with an evaporator and condensorrespectively connected to the inlet and outlet of the hermeticallysealed compressor unit of FIG. 1.

Referring in more detail to the accompanying drawings, FIG. 1illustrates a hermetically sealed compressor unit 10 embodying apparatusfor performingthe method of this invention in a vapor compressionrefrigeration system. Hermetic compressor unit 10 has an outer shell 12encasing a compressor 14 resiliently suspended therein which is drivenby an electric motor 16. Compressor I4 is of the positive displacementreciprocating piston type with a discharge leaf valve 18 mounted in adischarge chamber 19 of a cylinder head 20 secured to a cylinder block22 by bolts 24. Preferably, compressor 14 is a low-side casing type andhence has a discharge outlet coupling 26 fixed to a side wall of shell12 which is coupled with the interior of discharge chamber 19 throughtubular conduits 28 and 30. A refrigerant such as R-l2(dichlorodifluoromethane) is received in the casing space defined byhermetic shell 12 through an inlet coupling 32, and the bottom portionof the shell provides a reservoir for lubricating oil.

As shown in FIGS. 1 and 4, the refrigerant gas to be compressed issupplied from the interior of shell 12 to inlet valve ports 36 ofcompressor 14 through a pair of upright intake tubes 38 which aremounted at their lower ends in cylinder block 22 and each connected viaa muffler chamber 40, passageway 41, passageway 42 and suction chamber44 with ports 36 in cylinder head 20. Intake mufflers 40 and passages 41and 42 may be cored or machined in cast iron cylinder block 22. Eachintake tube 38 is press fit and/or silver soldered or adhesively securedin a counter bore 46 in the upper end of passageway 41. The compressorconstruction described above may be conventional and hence these andother details thereof will be well understood by those skilled in theart.

In accordance with a principal feature of the present invention, acalibrated orifice is selected and installed in the inlet passagewayleading to ports 36. In the example illustrated herein, this orifice ispreferably provided in the form of hollow restrictor plugs 50 which arereceived and fixed one in each of the upper ends of intake tubes 38, asby press fitting the plugs therein. As shown in FIGS. 2 and 3, each plug50 is generally cylindrical and has a shoulder 52 adjacent one end whichlimits the extent to which the plug can be inserted into the associatedtube 38. Each plug 50 has a fixed flow restriction in the form of acoaxial restricted orifice 54 therethrough which has a minimumcross-sectional area of less than one third and preferably in the rangeof one fourth to one tenth of the cross-sectional area of the insidediameter of its associated intake tube 38. Although the orifice 54 isfixed in size, it nevertheless provides a non-linear resistance to flowof gas therethrough; i.e., the pressure drop across orifice 54 variesdirectly with the mass flow rate of refrigerant gas therethrough inaccordance with the known fluid dynamics of restrictive orifices. Thesun of the volumes of both intake tubes 38, both passages 41, bothmufflers 40, both passages 42 and the common suction chamber 44 ispreferably in the range of two-to-five times the volumetric displacementof reciprocating piston compressor 14.

With respect to the location of the calibrated orifice, it is desirablethat there be a chamber of some given volume between the orifice and thesuction valve leaf. In the illustrated design, this chamber consists ofthe usual suction muffler chamber 40 and associated passages 41, 42 andthe suction chamber 44 formed in the head of the compressor plus theinterior of the suction tube 38 itself downstream from the restrictionplug 50. The optimum point at which plug 50 should be located from thestandpoint of the best starting characteristics is at the entrance tothe suction tube. However, in some installations from the standpoint ofnoise reduction, it may be better to install plug 50 at the downstreamend of suction tube 38 where it enters the muffler chamber proper. Ineither location the restricted orifice plug 50 in conjunction with thesuction tube 38 actually adds what amounts to additional muffling to theexisting muffler 40 and in this manner contributes to noise reduction.Orifice 54 may also be located at the point where the return line fromthe refrigeration system enters the compressor casing; i.e., in coupling32, rather than in the suction tube. However, this latter variationwould not give the noise reduction properties of the illustrated design.

As shown in FIG. 5, compressor unit with restricted orifices 54installed in the intake tubes thereof is used in a vapor compressorrefrigeration system with a condenser 56, capillary tube 58 and anevaporator 60 adapted for removing heat from a refrigerator cabinet 62.Discharge outlet coupling 26 is connected to the inlet of condenser 56and the outlet of the condenser is connected to the inlet of capillarytube 58. The outlet of capillary tube 58 is connected to the inlet ofevaporator 60 and the outlet of evaporator 60 is connected to inletcoupling 32 so that the compressor unit can circulate a refrigerantthrough the system.

In most automatic defrosting refrigerators, evaporator 60 is locateddirectly in refrigeration cabinet 62, and a thermostat senses thetemperature of the evaporator and cycles the compressor. In normaloperation of such a refrigerator, after the compressor has been shut offby the thermostat, the thermostat prevents the compressor from beingrestarted until the temperature of the evaporator in the foodcompartment rises to approximately 4045F. This rise in temperature meltsany ice formed on evaporator 60, thereby defrosting the evaporator. Inaddition, this entails sufficient down time to allow the refrigerationsystem pressures to equalize after every compressor pumping cycle.

In the so-called frost-free refrigerator, evaporator 60 is locatedoutside cabinet 62 and a fan circulates air over the evaporator and intothe cabinet to remove heat therefrom and cool the contents thereof. In afrost-free refrigerator defrosting cycle, hermetic compressor unit 10and the fan are shut off and an electric coil is used to heat theevaporator to approximately 65F thereby melting any ice formed thereonand also usually allowing sufficient time for the refrigeration systempressures to equalize. In some frost-free refrigerators, there aretypically four defrost cycles per day controlled by a timing clock. Inother frost-free refrigerators, there is a defrosting cycle every timethe compressor stops, which usually occurs two to four times per hour.

Also, the entire refrigeration system for all refrigerators andfreezers, regardless of their type and whether or not they have adefrost or frost-free cycle, usually reaches room or ambient temperaturewhenever the motor of the compressor unit is disconnected from itssource of power for any substantial period of time. Thus, there are manyoccasions when the evaporator of the refrigeration system becomes heatedsubstantially above its normal operating temperature and therefrigeration system reaches temperature as well as pressure equilibriumfor one reason or another.

In normal substantially steady state operation of the refrigerationsystem with a refrigerant such as R-l2, the compressor unit suctioninlet pressure at coupling 32 is substantially equal to the dischargepressure of evaporator 60 and is in the neighborhood of 20 p.s.i.a.Under such conditions, the discharge pressure of the compressor unit atcoupling 26 is substantially equal to the inlet pressure to condenser 56and is in the neighborhood of 200 p.s.i.a. if the refrigeration systemis cycled so that the evaporator temperature is maintained in the rangeof-lOF. to 0F. However, as previously indicated, whenever the evaporatoris heated to an abnormally high temperature, such as by defrosting or bydisconnecting the source of power to the compressor motor 16, therefrigeration system pressure rises abnormally and with the compressorshut down, pressure equalization occurs so that the refrigerantthroughout the system tends to equalize at pressures as high as 55p.s.i.a. when the temperature of the evaporator reaches approximately45F., approximately p.s.i.a. at approximately 65F and approximatelyp.s.i.a. at approximately 80F. As a result of this pressure rise, thedensity or mass per unit of volume of the gaseous refrigerant withinshell 12 of hermetic compressor 10 is I the compressor must work, thussubstantially increasing Jhe torque demand on the driving motor to 30seconds after the compressor is started.

However, due to the presence of orifices 54, when the compressor unit 10is restarted after the defrosting or other heating of evaporator 60, therate of flow of this high density gaseous refrigerant into compressor 14is automatically restricted, thereby limiting the amount of work done bythe compressor and thus, as previously indicated, reducing the maximumtorque load demand on the motor driving the compressor during theinitial start-up of the refrigeration system after the evaporator hasbeen heated. Usually in about 2 minutes after such a start-up thecompressor suction pressure at coupling 32 will have dropped to about 30p.s.i.a. and in about five minutes after start-up the discharge pressureat coupling 26 of the compressor will have dropped to about 200 p.s.i.a.Once the pressure at the compressor inlet has thus been reduced,orifices 54 due to their non-linear resistance to gas flow present adisproportionately lower resistance to gas flow therethrough and hencepresent only a slight increase in the pressure drop thereacross comparedto that existing through tubes 38 not equipped with restrictor plugs 50.

The invention is particularly well suited for use with low and mediumback pressure refrigeration systems wherein the density of therefrigerant increases sufficiently due to the heating of the evaporatorto provide an increase in the mass flow rate through the compressor ofsufficient magnitude that the power input or torque required to startand drive .the compressor is sharply increased. The back pressure, i.e.,the pressure of the refrigerant at the outlet of the evaporator andinlet to the compressor, is dependent on the particular refrigerant usedin the system and the temperature of the evaporator during normalsubstantially steady state operation of the refrigeration system. .Inlow back pressure systems, the pressure of the refrigerant at theevaporator outlet is generally in the range of 10 to 30 p.s.i.a. and inmedium back pressure systems, the pressure of the refrigerant at theevaporator outlet is in the range of to 45 p.s.i.a. for R-l2 and to 70p.s.i.a. for R-22. Low back pressure systems using a refrigerant such asR'l2 with the evaporator normally operat ing in the range of 40F. to10F. are commonly used in home refrigerators and freezers. Similarly,medium 7 back pressure systems using a refrigerant such as R-l2 or R-22with the evaporator normally operating in the range of l0F. to 30F. areused in package or beverage dispensers and commercial refrigerators suchas display cases.

The minimum surface area in radial cross section of orifice 54 throughplug 50 is selected to provide an optimum balance between theconflicting parameters of maximum output capacity and operatingefficiency of the hermetic motor compressor unit 10 during normaloperation thereof versus limiting the maximum torque required to run thecompressor under the abnormally large mass flow rate conditions. Undernormal running conditions the effect of the orifices as a resistance toquired to run the compressor under peakload conditions. This reductionin maximum torque imposed by abnormal load conditions in turn permits asmaller size motor to be used, or for the same size motor it provides ahigher rating for the compressor because of its ability to run undermore severe pressure conditions. In general, the smaller the minimumcross-sectionalsurface area of the orifices 54, the greater thereduction in the amount of peak torque or power required to run thecompressor and the lesser the overall efficiency and output capacity ofthe hermetic motor compressor unit. However, the cost benefits achievedfrom the reduction in the peak torque required to run a compressor inlow and medium back pressure refrigeration systems achieved by using acalibrated orifice pursuant to the present invention more than offsetsthe slight loss in overall efficiency, which may be in the range ofonehalf to 2 percent. If desired, the loss in capacity result ing fromthe installation of a calibrated orifice or orifices 54 can be readilycompensated for in existing compressor designs by increasing the borediameter of the cylinder or cylinders in compressor 14 to therebyincrease the displacement of the compressor an amount sufficient tooffset the reduction in pumping capacity caused by the orifices. Thischange can be accomplished at very little cost.

Most refrigeration cabinets have an optimum performance obtained byusing a compressor of a precise capacity rating, but no one modelcompressor may have this rating. Hence, compressors are quite oftenmismatched to some extent with the associated refrigeration systemequipment. Moreover, using an oversize or unduly high capacitycompressor with a given cabinet does not necessarilyproduce acorresponding increase in the output or performance of the cabinet inits refrigeration capacity. In accordance with the method of theinvention, an appropriate orifice plug or plugs 50 are selected for agiven compressor, thereby enabling a standardized compressor design tobe readily matched to varying types of refrigeration equipment.

Thus far it hasbeen found that a restriction in the form of the orifice54 provides better: characteristics for optimizing of peak torque versusrunning capacity as compared to a restriction in the form of a longnarrow tube. The calibrated orifice could be a plug 50 as shown or itcould be a thin plate or a mere necking down of the suction tube 38 toprovide the appropriate ratio of area change. Thus, in practicing themethod, the compressor manufacturer may provide a standardizedcompressor design capable of handling a range of system capacities.Then, in order to match this standard compressor to a given systemhaving a given capacity requirement, the maximum load conditionspresented by said system to said standard compressor are determinedwhereupon a calibrated orifice 54 may be installed in the compressoraccording to parameters disclosed herein to prevent overloading thereofto thereby match the compressor to said selected system.

Calibrated orifices 54 having a minimum crosssectional area providing apressure drop across the orifices in the range of 2 to 6 percent, andpreferably approximately four percent, of the absolute pressure of therefrigerant in shell 12 during normal substantially steady stateoperation of the refrigeration system (i.e., after the effects of theheated evaporator and pressure equalization have been dissipated) arebelieved to function satisfactorily. For the low back pressurerefrigeration system of FIG. 5, a pressure drop in the range of 0.4 to1.2 p.s.i., and preferably 0.8 p.s.i., is believed to be satisfactory.More particularly, for calculating an orifice such as passage 54 with acircular cross section, the diameter in inches of the minimum surfacearea in cross section should preferably be substantially equal to whereA is an empirical constant equal to 0.24, B is an empirical constantequal to 0.12, K is an empirical constant equal to 267, V is thespecific volume in cubic feet per pound mass of the gaseous refrigerantin shell 12 of the hermetic compressor during normal steady stateoperation of the refrigeration system, P is the desired pressure dropacross the orifice in pounds per square inch (it being understood thatthe value (P) is raised to the power expressed in the large brackets ofthe foregoing expression), and Q is the mass flow rate in pounds perhour through the orifice during normal substantially steady stateoperation of the refrigeration system.

With the use of calibrated orifices 54 in hermetic compressor unit 10,compressor 14 can be and preferably is driven by an electric motor 16incapable of producing sufficient maximum torque to initially start andrun the compressor after the evaporator has been heated and therefrigeration system pressure equalized in the absence of orifices 54.For example, it has been found that a hermeticcompressor unitconstructed in accordance with this invention having a single cylinderpositive displacement reciprocating piston-type compressor with avolumetric displacement of 1.067 cubic inches can be successfullystarted and run in a pressure equalized refrigeration system with aheated evaporator having an R-l 2 refrigerant by an electric motorcapable of developing a maximum torque of approximately 34 ounce feet atl volts when calibrated orifices 54 are utilized. Under the sameoperating conditions, the same electric motor will stall within 30seconds after the initial start-up of the compressor in the samerefrigeration system with the orifices 54 removed. This particularhermetic compressor unit had orifices providing a pressure drop of about4 percent of the normal substantialy steady state pressure of therefrigerant within shell 12 which resulted in a reduction of more than25 percent in the maximum torque required to start and run thecompressor in a pressure equalized system with a heated evaporator. Thissubstantial reduction in peak torque was achieved with a decrease ofless than 4 percent in the overall efficiency of the hermetic compressorunit 10 and a decrease of less than 10 percent of the maximum output ofthe compressor unit under the standard operating conditions specified inSection 6.2 of Standard 520 of the Air Conditioning and RefrigerationInstitute as published in 1968, with the evaporator at l0F., the gasentering the compressor at F., the compressor ambient temperature at 90Fthe liquid at the expansion valve at 90F. and the condensing temperatureat 130F. These tests were conducted in accordance with the Methods ofTesting for Rating Positive Displacement Refrigerant Compressorseffective June 25, 1967, of the American Society of Heating,Refrigeration and Air Conditioning Enginee rs, Inc. In this compressorunit there were two orifices 54 each having a minimum cross-sectionaldiameter of0. 125 inch in accordance with the above formula and. alength of three-eighths inch. The pressure drop across each orifice was0.8 pound per square inch when the upstream pressure was 19 p.s.i.a. andthe mass flow rate through the compressor during substantially steadystate operation was 24 pounds per hour. The crosssectional area of theopening of each orifice was approximately one-ninth of thecross-sectional area of its associated intake tubes and the combinedtotal volume of the intake tubes 38, intake passageways 41 and 42,intake mufflers 40 and suction chamber 44 was approximately seven cubicinches. The refrigeration system was cycled to maintain the temperatureof the evaporator in the range of l0F. to OF., except during defrostingwhen the evaporator was heated to raise the suction pressure top.s.i.a., and the compressor started and ran with l 10 volts to themotor.

Another example ofa hermetic compressor unit constructed and used in avapor compression refrigeration system in accordance with this inventionis a hermetic compressor unit having a positive displacementreciprocating piston-type compressor with a volumetric displacement of1.067 cubic inches, a motor developing a miximum torque of approximately34 ounce feet at 1 15 volts and two orifices 54 each having a diameterof .109 inch and a length of three-eighths inch. The pres sure dropacross each of the orifices 54 was approximately 1.0 p.s.i. when theupstream pressure was 19 p.s.i.a. The minimum cross-sectional area ofeach restricted orifice was approximately one-twelfth of its associatedintake tube and the combined total volume of the intake tubes 38, intakepassages 41 and 42, suction mufflers 40 and suction chamber 44 was 7cubic inches. This hermetic compressor unit was operated in arefrigeration system with an R-l2 refrigerant having a normalsubstantially steady state mass flow rate of 22.8 pounds per hourthrough the compressor with a normal substantially steady state pressureof 19 p.s.i.a. in shell 12 with the hermetic compressor unit beinginterniittently operated to maintain the temperature of the evaporatorbetween 10F. to 0F. except during defrosting when the evaporator washeated to raise the suction pressure to 104 p.s.i.a., and the compressorstarted and ran with volts to the motor.

A further example of a hermetic compressor unit constructed and used ina vapor compression refrigeration system in accordance with thisinvention is a hermetic compressor unit having a positive displacementreciprocating piston-type compressor with a volumetric displacement of1.067 cubic inches, a motor developing a maximum torque of approximately34 ounce feet at l 15 volts and two restricted orifices 54 each having adiameter of 0.140 inch and a length of three-eighths inch. The pressuredrop across each orifice was approximately 0.06 p.s.i. when the upstreampressure was 19 p.s.i.a. The minimum cross-sectional area of eachorifice 54 was approximately one-seventh of its associated intake tube-38 and the combined total volume of the intake tubes 38, intake passages41 and 42, suction mufflers 40 and suction chamber 44 was 7 cubicinches. This hermetic compressor unit was operated in a refrigerationsystem with an R-l2 refrigerant having a normal substantially steadystate mass flow rate of 24.3 pounds per hour through the compressor witha normal substantially steady state refrigerant pressure of 19 p.s.i.a.in shell 12 with the hermetic compressor unit being intermittentlyoperated to maintain the temperature of the evaporator between --lF. and0F. except during defrosting when the evaporator was heated to raise thesuction pressure to 87 p.s.i.a., and the compressor started and ran with1 l0 volts to the motor.

Both of these latter hermetic unit examples functioned' in arefrigeration system in generally the same manner and mode and producedthe same results as the first example described above.

By selecting and installing a restricted orifice plug 50 between thedischarge of an evaoprator and the inlet of the compressor in accordancewith the present invention, a vapor compression refrigeration apparatusand process is provided which substantially decreases the peak power ortorque requirements of the compressor in a pressure and/or temperatureequalized refrigeration system, particularly with a heated evaporator.The resulting hermetic compressor unit utilizes an electric drive motorhaving a substantially smaller maximum running torque capacity. Theinvention thus provides a hermetic compressor unit of more economicalconstruction and assembly compared to prior art hermetic compressorunits. Moreover, the use of a calibrated orifice with a fixed minimumcross-sectional area in the form of a simple, reliable and easilyinstalled orifice plug 50 enables this result to be accomplished with aninexpensive structure which is substantially service and maintenancefree throughout the useful life of the refrigeration system.

From the foregoing description, it will also now be apparent that thepresent invention is particularly useful in applications involving lowto medium back pressure refrigeration systems, although it is notnecessarily limited to such systems. Normally in such systems theinitial starting torque demand is less than the peak torque encounteredby motor 16 after the compressor 14 has been brought up to runningspeed, which may be for example 3,400 rpm. It only requires a fractionof a second, say one-tenth to one-half second, for the motor toaccelerate the compressor from zero velocity to its normal runningspeed. However, when the system is started after being shut down for asufficient period of time to cause partial or complete pressureequalization, which normally will occur anywhere from to 15 minutesafter shutdown, peak torque demand will be encountered normally fromabout to 30 seconds to about 1 minute after start-up of the compressor.This peak torque will normally last for a few minutes until the suctionor back pressure (normally measured in the compressor housing 12 priorto entry of the refrigerant gas into the suction tubes 38) has beenreduced approximately to normal operating values in the system.

Thus, in low to medium back pressure systems, the present inventionenables a compressor motor of reduced capacity to drive through thisperiod of peak loading by limiting the maximum gas pumping load whichcan be presented by the system to the compressor.

As indicated previously, the peak load conditions presented by varioustypes of vapor compression refrigeration systems will, of course, varydue to such factors as the system having a defrost cycle or frost-freemode of operation, as well as the type of condition encountered atstart-up of the compressor. During the occurrence of a soak-outcondition (i.e., wherein the system is shut off for an extended periodof time, normally 18 to 24 hours, so that both temperature and pressureequalization will occur throughout the system), an abnormal gas pumpingload may be imposed on the compressor. However, pressure equalizationcan also occur without temperature equilibrium having been established,as in the case of a defrost cycle. At the start of such a cycle, thecondenser is already hot, the condenser gas temperature being say in therange of F. to F., and then due to warming of the evaporator the gastemperature therein will be faised from say 0F. to 40F. This presents anaggravated load situation upon restarting of the compressor as comparedto a soak-out condition wherein system temperature may have equalized ata lower temperature, say anywhere from 60F. to 1 10F, depending uponambient temperature conditions. The peak load problems, of course, areencountered at the high end of such temperature ranges. Normally thereis not enough gas in the system to create enough pressure to causecondensation of the gas in the condenser at such elevated temperatures.Hence, saturated vapor conditions may not prevail and, if so, thepressures do not follow the saturated vapor curves.

Moreover, during a soak-out condition, the lubricating oil present inthe oil sump at the bottom of shell or housing 12 will drop intemperature. This factor, coupled:with the quiescent state of the oil,will allow the oil to absorb some of the refrigerant vapor. When thevapor goes into solution in the oil, the system will be depleted of someof its refrigerant in vapor form, thereby reducing the gas pumping loadencountered in pul ing down the system after start-up. However, during ashorter shutdown condition, such as occurs in a defrost cycle, the oilremains too hot to absorb much of thie refrigerant vapor and hence morerefrigerant in vapogr form will be present in the system. Moreover, thegas pressure can reach higher levels because the motor will "still behot from its previous operating cycle, and there is more gas in thesystem because less of the gas can be absorbed by the higher temperatureoil. Also, the condenser will be hotter than the ambient temperature andhence the discharge pressure of the compressor will be higher afterstart-up and the rise in discharge pressure will be accelerated.Nevertheless, if the gas temperature is high enough in the system, therestill may not be enough refrigerant present to produce saturated gasconditions. Hence, short shutdown cycles may pose more severe gasloading problems than encountered after a soak-out condition hasoccurred.

Due to the various factors mentioned previously, the density of therefrigerant gas at the inlet to the compressor may be at least doubledcompared to the density of the refrigerant gas during normal operationof the refrigeration system. In many instances there will be atwo-to-five fold increase in the density of the gaseous refrigerant whenthe evaporator is heated from say F. to about 80F., assuming a saturatedvapor condition. However, after the evaporator temperature reaches about80F., usually no liquid refrigerant will be left in the system. Hence,saturated vapor conditions will no longer prevail and, therefore, abovethis temperature the pressure does not increase as fast due to thepresence of superheated vapor conditions in the system. Nevertheless,until such superheated vapor conditions are reached, the aforementionedtwo-to-five fold density change or increase in back pressure can anddoes occur.

The aforementioned peak gas pumping load problems normally may not be asserious in high back pressure systems but the concepts of the presentinvention can also be utilized to advantage in such systems. Byproviding a suitably calibrated orifice means in such a system, gasloading imposed on the gas pump of the compressor can be limited,thereby enabling the amount of refrigerant charge in the system to beincreased. In other words, by suitably throttling gas flow into thecompressor housing in such a system, the system can tolerate a greateramount of refrigerant charge before the problems of oil pump out orliquid entering the cylinder of the compressor will be encountered.Also, in high back pressure systems, the provision of a calibratedorifice means of the present invention will reduce the maximum pressureand temperature conditions produced in the system when the condenser fanfails or the air flow through the condenser is reduced or restricted dueto other adverse conditions, such as dirt or dust accumulation in thecondenser, or when the ambient temperature condition of the condenser isabnormally high.

I claim:

1. A method of decreasing the peak torque demand of a hermeticallysealed positive displacement compressor in a vapor compressionrefrigeration system containing a refrigerant and wherein a low ormedium back ressure exists at the outlet of an evaporator of the systemduring normal operation thereof, said method comprising the steps ofproviding a compressor having a volumetric displacement rating in excessof the requirements of said system, selecting flow restriction orificemeans having a fixed minimum cross-sectional area calibrated to limitthe mass flow rate of said refrigerant into said compressor after it isjust started in the operation of said system under an abnormal conditionwith said evaporator at an elevated temperature relative to its normaloperating temperature, and locating said orifice means in saidrefrigeration system such that all of the refrigerant entering thesuction inlet of the compressor does so exclusively via said orificemeans, whereby the peak torque required to run the compressor under saidabnormal condition is decreased compared to the torque required to runsaid compressor under the same conditions without said orifice means.

2. The method of claim 1 in which said orifice means is located betweenthe outlet of the evaporator and the suction inlet of the compressor.

3. The method of claim 2 in which said orifice means is located in apassageway and said minimum area of said flow restriction means isdimensioned to be less than one-third of the minimum cross-sectionalarea of the passageway immediately downstream of said flow restrictionmeans of said orifice means.

4. The method of claim 2 in which said restricted orifice means islocated in a passageway and said minimum area of said flow restrictionmeans is dimensioned to be in the range of one-third to one-tenth of theminimum area of the passageway immediately downstream of said flowrestriction means.

5. The method of claim 2 in which the fixed minimum cross-sectional areaof said flow restriction means is selected such that the averagepressure drop thereacross is less than 6 percent of the normal operatingpressure of the refrigeration system at the outlet of the evaporator.

6. The method of claim 2 in which the fixed minimum cross-sectional areaof said flow restriction means is selected such that the averagepressure drop thereacross is less than 1 pound per square inch absoluteduring normal substantially steady state operation of the refrigerationsystem.

7. The method of claim 2 in which said flow restriction means isselected such that the average pressure drop thereacross is in the rangeof four-tenths to one and two-tenths pounds per square inch duringnormal substantially steady state operation of the refrigeration system.

8. The method of claim 2 in which said flow restriction means isselected such that the decrease in overall efficiency of the hermeticcompressor due to coupling the outlet of the evaporator with the suctioninlet of the compressor through the flow restriction means is less thanfive percent.

9. The method of claim 2 in which the fixed minimum cross-sectional areaof said flow restriction means is generally circular and has a diameterD in inches substantially equal to where K equals 167, A equals 0.24, 8equals 0.12, V equals the specific volume in cubic feet per pound massof the refrigerant immediately upstream of the orifice, P equals thepressure drop across the flow restriction means in pounds per squareinch, and Q equals the mass flow rate in pounds per hour through thecompressor when the refrigeration system is operating under theconditions of Section 6-2 of Standard 520 of the American RefrigerationInstitute, with the evaporator temperature at about --l0F., the gastemperature entering the compressor at about F., the compressor ambienttemperature at about 90F., the liquid temperature at the expansion valveat about 90F., and the condensing temperature at about F.

10. The method of claim 9 in which said flow restriction means islocated in an inlet tube coupled to a muffler of the hermetically sealedcompressor.

11. The method of claim 2 in which said flow restriction means islocated in an inlet tube coupled to a muffler of the hermetically sealedcompressor.

12. The method of claim 1 wherein an electric motor is provided to drivethe compressor and said motor is selected so as to be incapable ofproducing sufficient torque to start and run the compressor when saidrefrigeration system is in a temperature and/or pressure equalizedcondition with said evaporator at an elevated temperature without saidrestricted orifice.

13. In a hermetic compressor unit for a refrigeration system having ahermetic casing with inlet means for supplying a refrigerant gas to thespace in said casing, a positive displacement compressor with an inletvalve, said compressor being mounted in said casing and adapted to bedriven by an electric motor, a muffier chamber having an outletcommunicating with said compressor inlet valve and a suction pipeextending generally upright in said casing and communicating at itslower end with an inlet of said muffler chamber and at its upper endwith the space in said casing, the improvement comprising restrictedorifice means having a fixed minimum cross-sectional area in saidsuction pipe for restricting the mass flow rate of refrigeranttherethrough, said orifice means being located such that all of therefrigerant enters said compressor exclusively via said restrictedorifice means and being dimensioned to limit the mass flow rate ofrefrigerant to a given value through said compressor when saidcompressor is started under a given condition of abnormally high densityof gaseous refrigerant in said casing of at least twice the density ofthe gaseous refrigerant in said casing during normal operation of thereferigeration system, and an electric motor received in said casing andadapted to drive said compressor, said electric motor being incapable ofproducing the peak torque imposed by the gas pumping load exerted onsaid compressor under said given condition in the absence of saidorifice means and being capable of developing the maximum torque imposedby the gas pumping load exerted on said compressor under said givencondition with said orifice means in said suction pipe.

14. The hermetic compressor unit of claim 13 wherein said restrictedorifice means is located at the upper .end of said suction pipe.

15. The hermetic compressor unit of claim 13 wherein said restrictedorifice means is located in the vicinity of the connection of saidsuction pipe with said muffler chamber.

' 16. The hermetic compressor unit of claim 13 wherein the combinedtotal volume of all passageways and chambers between said restrictedorifice means and said inlet valve of said gas compressor is in therange of two to five times the volumetric displacement of saidcompressor.

17. The hermetic compressor unit of claim 13 in which the minimumcross-sectional area of said restricted orifice means is less thanone-third of the minimum cross-sectional area of the passageway definedby said suction tube downstream of said restricted orifice means.

18. The hermetic compressor unit of claim 17 in which said minimumcross-sectional area of said restricted orifice means is such that theaverage pressure drop across said restricted orifice means is less than6 percent of the pressure of the refrigerant gas in said hermetic casingduring normal substantially steady state operation of the refrigerationsystem in which the compressor unit is utilized.

19. The hermetic compressor unit of claim 17 in which said minimumcross-sectional area of said restricted orifice means is such that theaverage pressure drop across the restricted orifice means is in therange of 0.4 to 1.2 pounds per square inch during normal substantiallysteady state operation of the refrigeration system in which thecompressor unit is utilized.

20. The hermetic compressor unit of claim 17 in which said minimumcross-sectional area of said restricted orifice means is such that theoverall efficiency of the hermetic compressor unit during substantiallysteady state operation in the refrigeration system is decreased lessthan 5 percent due to said restricted orifice means in said suctionpipe.

21. The hermetic compressor unit of claim 17 in which said minimumcross-sectional area of said restricted orifice means is generallycircular and has an effective diameter D in inches substantially equalto I BP K(P) FPE where K equals 267, A equals 0.24, B. equals 0.12, Vequals the specific volume in cubic feet per pound mass of therefrigerant immediately upstream of the orifice, P equals the pressuredrop across the orifice in pounds per square inch, and Q equals the massflow rate through the compressor in pounds per hour where therefrigeration system is operating under the conditions of Section 6-2 ofStandard 520 of the American Refrigeration Institute, with evaporatortemperature at l0F., gas temperature entering the compressor at about Fcompressor ambient temperature at about 90F, liquid temperature at theexpansion valve at about 90F., and condensing temperature at about F.

22. A method of decreasing the peak gas pumping torque imposed on ahermetically sealed positive displacement compressor to match thecompressor with a vapor compressor refrigeration system having acondenser, expansion valve means and an evaporator serially connectedwith said compressor and containing a refrigerant, and wherein a low tomedium back pressure exists at the outlet of the evaporator of thesystem during normal operation thereof, said method comprising the stepsof providing a compressor having a volumetric displacement rating inexcess of that required by said system, selecting restricted orificemeans calibrated to reduce the pumping capacity of said compressor tomatch the requirements of said system with said compressor, and locatingsaid orifice means in said system such that the outlet of the evaporatorcommunicates with the suction inlet of the compressor through saidorifice means whereby said compressor is matched with said refrigerationsystem.

23. The method of claim 22 wherein said orifice means has a fixedminimum cross-sectional area and is located between the outlet of theevaporator and the suction inlet of the compressor so as to continuouslycouple the same through said orifice means.

24. A vapor compression refrigeration system containing a refrigerantand having a low to minimum back pressure at the outlet of an evaporatorof the system during normal operation of said system, said systemfurther having the inlet of a hermetically sealed positive displacementcompressor coupled to the outlet of the evaporator exclusively viaorifice means having a fixed minimum cross-sectional area calibrated tolimit the maximum mass flow rate into the compressor after it is firststarted and an abnormal back pressure occurs with the evaporator at anelevated temperature compared to its normal operating temperaturecompared to its normal operating temperature whereby the torque requiredto operate the compressor is decreased compared to the torque requiredto operate the compressor under the same conditions without said orificemeans.

25. The refrigeration system of claim 24 wherein said hermeticallysealed positive displacement compressor comprises an electric motor, apositive displacement gas pump driven by said motor, and a casingencapsulating and hermetically sealing said motor and said gas pumptherein, said electric motor being incapable of producing sufficienttorque to operate said gas pump with said evaporator at said elevatedtemperature in the absence of said orifice means and being capable ofdeveloping sufficient torque to operate said gas pump with said orificemeans.

26. The refrigeration system of claim 24 wherein said elevatedtemperature of said evaporator is sufciciently greater than said normaloperating temperature to at least double the density of the gaseousrefrigerant at the outlet of said evaporator compared to the density ofthe gaseous refrigerant at said outlet of said evaporator at said normaloperating temperature.

27. The refrigeration system of claim 25 wherein said elevatedtemperature of said evaporator is sufficiently greater than said normaloperating temperature to at least double the density of the gaseousrefrigerant at the outlet of said evaporator compared to the density ofthe gaseous refrigerant at said outlet of said evaporator at said normaloperating temperature.

7 3 UNITED STATES PATENT OFFICE a s9 w r CERTIFICATE OF CORBLCHON PatentNO. 3, 763,659 Dated October 9, 1973 Paul B. Hover Invcnt0r(s) It iscertified that error appears in the above-identified patent and thatsaid Letters Patent are hereby corrected as shown below:

Column 5, line 11, "Jhe" should be -the--.

In the formula appearing in columns 7, l2 and 14, that portion thereofreading "V0" should read VQ--.

Column 12, line 47, "167" should be 267- Column 13, line 21, cancelpipe' and insert tube means;

line 42, cancel "pipe" and insert tube means-; line 45, cancel "pipe"and insert tube means-; line 57, after "tube" insert Column 14, line 14,cancel "pipe" and insert tube means-L Signed and sealed this 25th day ofJune l97L| (SEAL) Attest:

EDWARD M.FLETCHE.R,JR. C. MARSHALL DANN Attesting Officer Commissionerof Patents L. a I r

1. A method of decreasing the peak torque demand of a hermeticallysealed positive displacement compressor in a vapor compressionrefrigeration system containing a refrigerant and wherein a low ormedium back pressure exists at the outlet of an evaporator of the systemduring normal operation thereof, said method comprising the steps ofproviding a compressor having a volumetric displacement rating in excessof the requirements of said system, selecting flow restriction orificemeans having a fixed minimum cross-sectional area calibrated to limitthe mass flow rate of said refrigerant into said compressor after it isjust started in the operation of said system under an abnormal conditionwith said evaporator at an elevated temperature relative to its normaloperating temperature, and locating said orifice means in saidrefrigeration system such that all of the refrigerant entering thesuction inlet of the compressor does so exclusively via said orificemeans, whereby the peak torque required to run the compressor under saidabnormal condition is decreased compared to the torque required to runsaid comprEssor under the same conditions without said orifice means. 2.The method of claim 1 in which said orifice means is located between theoutlet of the evaporator and the suction inlet of the compressor.
 3. Themethod of claim 2 in which said orifice means is located in a passagewayand said minimum area of said flow restriction means is dimensioned tobe less than one-third of the minimum cross-sectional area of thepassageway immediately downstream of said flow restriction means of saidorifice means.
 4. The method of claim 2 in which said restricted orificemeans is located in a passageway and said minimum area of said flowrestriction means is dimensioned to be in the range of one-third toone-tenth of the minimum area of the passageway immediately downstreamof said flow restriction means.
 5. The method of claim 2 in which thefixed minimum cross-sectional area of said flow restriction means isselected such that the average pressure drop thereacross is less than 6percent of the normal operating pressure of the refrigeration system atthe outlet of the evaporator.
 6. The method of claim 2 in which thefixed minimum cross-sectional area of said flow restriction means isselected such that the average pressure drop thereacross is less than 1pound per square inch absolute during normal substantially steady stateoperation of the refrigeration system.
 7. The method of claim 2 in whichsaid flow restriction means is selected such that the average pressuredrop thereacross is in the range of four-tenths to one and two-tenthspounds per square inch during normal substantially steady stateoperation of the refrigeration system.
 8. The method of claim 2 in whichsaid flow restriction means is selected such that the decrease inoverall efficiency of the hermetic compressor due to coupling the outletof the evaporator with the suction inlet of the compressor through theflow restriction means is less than five percent.
 9. The method of claim2 in which the fixed minimum cross-sectional area of said flowrestriction means is generally circular and has a diameter D in inchessubstantially equal to
 10. The method of claim 9 in which said flowrestriction means is located in an inlet tube coupled to a muffler ofthe hermetically sealed compressor.
 11. The method of claim 2 in whichsaid flow restriction means is located in an inlet tube coupled to amuffler of the hermetically sealed compressor.
 12. The method of claim 1wherein an electric motor is provided to drive the compressor and saidmotor is selected so as to be incapable of producing sufficient torqueto start and run the compressor when said refrigeration system is in atemperature and/or pressure equalized condition with said evaporator atan elevated temperature without said restricted orifice.
 13. In ahermetic compressor unit for a refrigeration system having a hermeticcasing with inlet means for supplying a refrigerant gas to the space insaid casing, a positive displacement compressor with an inlet valve,said compressor being mounted in said casing and adapted to be driven byan eleCtric motor, a muffler chamber having an outlet communicating withsaid compressor inlet valve and a suction tube means extending generallyupright in said casing and communicating at its lower end with an inletof said muffler chamber and at its upper end with the space in saidcasing, the improvement comprising restricted orifice means having afixed minimum cross-sectional area in said suction tube means forrestricting the mass flow rate of refrigerant therethrough, said orificemeans being located such that all of the refrigerant enters saidcompressor exclusively via said restricted orifice means and beingdimensioned to limit the mass flow rate of refrigerant to a given valuethrough said compressor when said compressor is started under a givencondition of abnormally high density of gaseous refrigerant in saidcasing during normal operation of the refrigeration system, and anelectric motor received in said casing and adapted to drive saidcompressor, said electric motor being incapable of producing the peaktorque imposed by the gas pumping load exerted on said compressor undersaid given condition in the absence of said orifice means and beingcapable of developing the maximum torque imposed by the gas pumping loadexerted on said compressor under said given condition with said orificemeans in said suction tube means.
 14. The hermetic compressor unit ofclaim 13 wherein said restricted orifice means is located at the upperend of said suction tube means.
 15. The hermetic compressor unit ofclaim 13 wherein said restricted orifice means is located in thevicinity of the connection of said suction tube means with said mufflerchamber.
 16. The hermetic compressor unit of claim 13 wherein thecombined total volume of all passageways and chambers between saidrestricted orifice means and said inlet valve of said gas compressor isin the range of two to five times the volumetric displacement of saidcompressor.
 17. The hermetic compressor unit of claim 13 in which theminimum cross-sectional area of said restricted orifice means is lessthan one-third of the minimum cross-sectional area of the passagewaydefined by said suction tube downstream of said restricted orificemeans.
 18. The hermetic compressor unit of claim 17 in which saidminimum cross-sectional area of said restricted orifice means is suchthat the average pressure drop across said restricted orifice means isless than 6 percent of the pressure of the refrigerant gas in saidhermetic casing during normal substantially steady state operation ofthe refrigeration system in which the compressor unit is utilized. 19.The hermetic compressor unit of claim 17 in which said minimumcross-sectional area of said restricted orifice means is such that theaverage pressure drop across the restricted orifice means is in therange of 0.4 to 1.2 pounds per square inch during normal substantiallysteady state operation of the refrigeration system in which thecompressor unit is utilized.
 20. The hermetic compressor unit of claim17 in which said minimum cross-sectional area of said restricted orificemeans is such that the overall efficiency of the hermetic compressorunit during substantially steady state operation in the refrigerationsystem is decreased less than 5 percent due to said restricted orificemeans in said suction tube means.
 21. The hermetic compressor unit ofclaim 17 in which said minimum cross-sectional area of said restrictedorifice means is generally circular and has an effective diameter D ininches substantially equal to
 22. A method of decreasing the peak gaspumping torque imposed on a hermetically sealed positive displacementcompressor to match the compressor with a vapor compressor refrigerationsystem having a condenser, expansion valve means and an evaporatorserially connected with said compressor and containing a refrigerant,and wherein a low to medium back pressure exists at the outlet of theevaporator of the system during normal operation thereof, said methodcomprising the steps of providing a compressor having a volumetricdisplacement rating in excess of that required by said system, selectingrestricted orifice means calibrated to reduce the pumping capacity ofsaid compressor to match the requirements of said system with saidcompressor, and locating said orifice means in said system such that theoutlet of the evaporator communicates with the suction inlet of thecompressor through said orifice means whereby said compressor is matchedwith said refrigeration system.
 23. The method of claim 22 wherein saidorifice means has a fixed minimum cross-sectional area and is locatedbetween the outlet of the evaporator and the suction inlet of thecompressor so as to continuously couple the same through said orificemeans.
 24. A vapor compression refrigeration system containing arefrigerant and having a low to minimum back pressure at the outlet ofan evaporator of the system during normal operation of said system, saidsystem further having the inlet of a hermetically sealed positivedisplacement compressor coupled to the outlet of the evaporatorexclusively via orifice means having a fixed minimum cross-sectionalarea calibrated to limit the maximum mass flow rate into the compressorafter it is first started and an abnormal back pressure occurs with theevaporator at an elevated temperature compared to its normal operatingtemperature compared to its normal operating temperature whereby thetorque required to operate the compressor is decreased compared to thetorque required to operate the compressor under the same conditionswithout said orifice means.
 25. The refrigeration system of claim 24wherein said hermetically sealed positive displacement compressorcomprises an electric motor, a positive displacement gas pump driven bysaid motor, and a casing encapsulating and hermetically sealing saidmotor and said gas pump therein, said electric motor being incapable ofproducing sufficient torque to operate said gas pump with saidevaporator at said elevated temperature in the absence of said orificemeans and being capable of developing sufficient torque to operate saidgas pump with said orifice means.
 26. The refrigeration system of claim24 wherein said elevated temperature of said evaporator is sufcicientlygreater than said normal operating temperature to at least double thedensity of the gaseous refrigerant at the outlet of said evaporatorcompared to the density of the gaseous refrigerant at said outlet ofsaid evaporator at said normal operating temperature.
 27. Therefrigeration system of claim 25 wherein said elevated temperature ofsaid evaporator is sufficiently greater than said normal operatingtemperature to at least double the density of the gaseous refrigerant atthe outlet of said evaporator compared to the density of the gaseousrefrigerant at said outlet of said evaporator at said normal operatingtemperature.